1. Field of the Invention
The present invention relates to a shape of a screw rotor for use in oil-cooled type screw compressor or expander and, more particularly, to a screw rotor having a shape suitable for hobbing.
2. Description of the Prior Art
Generally speaking, a screw rotor of asymmetric teeth type, used in screw compressors or the like, has a female rotor and a male rotor in a pair, the female rotor having its major part at the inside of the pitch circle, while the male rotor having its major part at the outside of the pitch circle.
FIG. 1 shows an example of the screw rotor of the kind described. A female rotor 1 and a male rotor 2 meshing with each other are adapted to rotate around respective axes 3, 4 within the casing (not shown) in the direction of arrows to function as a compressor. The female rotor 1 is provided with a multiplicity of grooves 5 and ridges 6. Each groove 5 has a leading side flank 7, a first trailing side flank 8, a second trailing side flank 9 and a teeth bottom flank 10 interconnecting the flanks 7 and 8. These flanks constitute a major part of the female rotor, the major part being disposed at the inside of a pitch circle 11. On the other hand, the male rotor 2 has a plurality of ridges 12 and grooves 13. Each ridge 12 has a major part constituted by a leading side flank 14, first trailing side flank 15, second trailing side flank 16 and a teeth end flank 17 interconnecting both flank 14 15. The major part constituted by these flanks is disposed at the outside of the pitch circle 18.
The portion between points 19 and 20 of the leading side flank 7 of the female rotor 1 has an arcuate form centered at a point 22 which is located on the extension of a line interconnecting the point 20 and the point of intersection of the pitch circles 11, 18, i.e. the pitch point 21, and is positioned such that the line interconnecting the points 19 and 22 is normal to the radial line at the point 19. The portion between points 23 and 24 of the first trailing side flank 8 is formed by a curve which is created by the junction point 25 between the first trailing side flank 15 and the teeth end flank 17 of the male rotor 2. The portion between points 24 and 26 on the second trailing side flank 9 is formed by the extension of a line interconnecting the center 3 of rotation and the point 24.
Referring now to the male rotor 2, the portion between points 27 and 28 of the leading side flank 14 is a curve created by the curve between the points 19 and 20 of the leading side flank of the female rotor 1. The portion between points 25 and 29 of the first trailing side flank 15 is a curve created by the point 24 of the female rotor 1. The portion between points 29 and 30 of the second trailing side flank 16 is a curve created by the portion between points 24 and 26 of the female rotor 1. Finally, the portion between points 28 and 25 of the teeth end flank 17 is an arc centered at the pitch point 21.
At the outside of the pitch circle 11 of the ridge 6 of the female rotor 1, are disposed the follow portion between points 19 and 31 and follow portion between points 26 and 32. These points 31, 32 are positioned at the top of respective ridges.
Similarly, at the inside of the pitch circle 18 of the groove 13 of male rotor 2, are disposed a follow portion between points 27 and 33 and a follow portion between points 30 and 34. The points 33 and 34 are positioned on the bottom of groove 13.
The screw rotor having the described shape is not suitable for hobbing.
Namely, a pressure angle is zero at each of the points 19, 26, 27, 30 located on the pitch circles 11, 18 of both rotors 1, 2, so that a generating pitch circle is set at the outside of the pitch circle of the rotor. If this generating pitch circle is made excessively large as compared with the pitch circle of the rotor, the minimum pressure angle on the hobbing edge is increased to advantageously prolong the life of the tool. This, however, on the other hand causes a large polygonal error, resulting in a deteriorated error of shape of the rotor. Therefore, the setting of the generating pitch circle is limited by both of the tool life and the precision of the rotor. In addition, in the cutting of this rotor, it is not possible to preserve a sufficiently large radius of curvature of the hobbing edge for male rotor, which shares the greatest cutting amount, so that the tool is rapidly worn out locally at this position. Furthermore, since the hobbing cutter edge has a small minimum pressure angle for its teeth height, it is difficult to machine the hobbing edge at a sufficiently high precision, resulting in a raised cost of the hob.
The conventional screw rotor also has the following disadvantage also in the aspect of performance.
Generally, the performance of a screw compressor incorporating screw rotors is influenced by various factors. As to the shape of the rotor, these factors are the length of the seal line and the area of the blow hole.
The length of the seal line is the contact length between the teeth of rotors, and the product of this contact length and the gap between rotors is the area of leakage. FIG. 2 shows the projection of the locus of contact point of screw rotors shown in FIG. 1, on a cross-section perpendicular to the axis.
Referring to FIG. 2, the locus of contact point between the follow portions of the leading sides of the rotors 1, 2 is shown at curve a-b-c. Similarly, the locus of contact points between follow portions of the trailing side is shown at curve a-d-c. The locus of contact point between the leading side flanks is shown at curve a-e. The locus of contact point between the teeth bottom flank of the female rotor 1 and the teeth end flank of the male rotor 2 is shown at a curve e-f. The locus of contact point between the first trailing side flanks is shown at a curve f-g and a curve g-h. The locus of contact point between the second trailing side flanks is shown at curve h-a. These screw rotors have large lengths of sealing lines at the locus of contact point a-b-c, a-d-c and a-e. The large length of the seal line between rotors causes a correspondingly increased area of leakage, resulting in a deteriorated performance of the compressor.
In FIG. 2, the blow hole formed between the edge point i of the high-pressure casing and the point g on the above-mentioned locus is large because the distance between the points i and g are large due to the fact that the portion between points 24 and 26 of the second trailing side flank 9 of the female rotor 1 is formed by a straight line and due to the pressure of the follow portions between points 26 and 32. The large blow hole has a tendency to permit the leakage of the fluid from the high-pressure chamber to the low-pressure chamber to deteriorate the performance of the compressor.
A screw rotor for overcoming the above-described problem is disclosed, for example, in the specification of U.S. Pat. No. 3,787,154. This screw rotor, however, cannot provide a satisfactory solution to the problem of the tool life, blow hole area and so forth.